Methods of compressing fluids with centripetal compressors



V. H. PAVLECKA June 26, 1962 METHODS OF' COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 1 w. m m m r rae/ffy.

June 26, 1962 v. H. PAvLEcKA 3,040,971

METHODS OF COMPRESSING' FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 2 4free/Vix June 26, 1962 v. H. PAvLEcKA METHODS OF1 COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS 14 Sheets-Sheec 3 Original Filed June 8, 1955 w @NN INVENToR. VZ //f//f/e if P41/raw #rra/ewig June 26, 1962 V. H. PAVLECKA METHODS OF COMPRESSING FLUIDS WITH CENTRIIPETU.. COMPRESSORS 14 Sheets-Sheet 4 Original Filed June 8, 1955 INVENToR. PZ ,4a/ffm@ HP/Q K4 aan @fw/77M June 26, 1962 v. H. PAvLEcKA METHODS OF COMPRESSING F'LUIDS WITH CENTRIPETAL COMPRESSORS 14 Sheets-Sheet 5 Original Filed June 8, 1955 June 26, 1962 v, H. PAVLECKA 3,040,971

METHODS OF' COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 6 E515. Ufa .1

/4 En. 12. 'INVENTOR 0* I VZa/M/,efawffm June 26, 1962 v. H. PAVLECKA 3,040,971

METHODS OF' COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Shee'c 7 INVENToR. 4 V2M/,me 15C P41/z am innen/5% June 26, 1962 v. H. PAvLl-:CKA

METHODS OF' COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 8 v @GM/ w wmwN h @Q u NQS Q NQWN June 26, 1962 v. H. PAVLECKA 3,040,971

METHODS OF COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 9 6T, E6. 2j.

2200v EG. 22.

a; Leal June 26, 1962 v. H. PAvLEcKA 3,040,971

METHODS OF COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 1G June 26, 1962 v. H. PAvLEcKA METHODS OF' COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 1l .Nm SE1 INVENTOR. VZ/m//w/e .zD V4 cfg BY H/lc 7 inve/uff! June 26, 1962 v. H. PAvLEcKA METHODS OF COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 12 Nm GNN wm SNN rraeA/[H June 26, 1962 v. H. PAvLEcKA 3,040,971

METHODS OF COMPRESSING FLUIDS WITH CENTRIPETAL COMPRESSORS Original Filed June 8, 1955 14 Sheets-Sheet 13 June 26, 1962 V. H. PAVLECKA METHODS OF' COMPRESSING FLUIDS WITH CENTRIPTAL COMPRESSORS 14 Sheets-Sheet' 1.4

Original Filed June 8, 1955 INVENTOR.

7T k rra/ew'fff United States Patent Oil ce This application for patent is a continuation of my earlier application, Serial Number 514,001, filed lune 8,

1955, entitled Methods of Compressing Fluids With- Centripetal Compressors, now abandoned, which is being replaced with this application.

This invention relates to novel methods of compressing elastic fluids by means of centripetal flow dynamic cornpressors.

It is an object of this invention -to provide novel cornpression methods for centripetal ilow compressors, in which loading of the stages is apportioned so as to operate as many stages as possible at constant local Mach numbers for obtaining maximum total head from a given number of compression stages.

Still another object of this invention is to provide novel compression methods for subsonic flow compressors in which energy contributions of the respective stages are increased by introducing one or several stationary vector-adjusting stages between the outermost compression stages for obtaining a more effective and `uniform distribution of the fluid dynamic energies among all compression stages.

The novel features which are believed to be characteristic of the invention, both as to its organization and method of operation, together with further objects and advantages thereof, will be better understood from the following description taken in connection with the accompanying drawings in which several embodiments of the invention are illustrated as examples of the invention. Referring to -the drawings:

FIG. 1 is a transverse sectional view of a centripetal being positioned between two contra-rotating compression stages.

FIG. 19 is an axial sectional view of a centripetal flow compressor with two vector-adjusting stages, both of the vector-adjusting stages being positioned between the corotating stages.

FIG. 20 is a diagram of peripheral velocities of uids through a centripetal ilow subsonic compressor shown in FIG. 19 by the resultant curve indicating: the changes in these velocities as the uid travels from an outer periphery of the compressor toward its axis of rotation.

FIG. 21 is the same curve as that illustrated in FIG. 20, but yfor a classical centripetal compressor having two contra-rotating rotors and a symmetrical vector diagram such as that illustrated in FIG. 25.

FIG. 22 is the same curve `for a comrpressor having one vector-adjusting stage positioned between contra-rotating stages, i.e., it is the peripheral velocities curve for a compresser similar to the compressor illustrated in FIGS. 7, 8 and 9, but having nine compression stages instead of 5.

FIG. 23 is the peripheral velocities curve for a compresser similar to the compressor illustrated in FIGS. 13,

` eral velocity vectors (Cu)s for a compressor having a compressor .with asymmetric arrangement of compression stages;

FIG. 2 is a vector diagram for the compressor illustrated in FIG. l;

FIGS. 3 and 4A through 4C are explanatory vector diagrams;

FIG. 5 is a diagram illustrating the chordal angle e.

FIGS. 6 and 9 are vector diagrams for the centripetal subsonic compressor illustrated in FIG. 7 which has a single vector-adjusting stage between the rst and the second compression stage;

FIGS. 7 and 8 are axial and transverse sections, respectively, of a centripetal compressor with a single vectoradusting stage and two contra-rotating rotors with the vector-adjusting stage being positioned between the rst and second contra-rotatable compression stages;

FIGS. l() and l1 are axial and transverse sections, respectively, of a centripetal subsonic compressor having two vector-adjusting stages, each vector stage being positioned between contra-rotating stages;

FIG. 12 is a vector diagram of the compressor of FIGS. 10 and 11;

FIGS. 13 and 14 are axial and transverse sectional views of a centripetal ow compressor having a single vector-adjusting stage positioned between two co-rotating compression stages and FIG. l5 is a vector diagram for this compressor;

FIGS. 16, 17 4and 18 are the sectional views and the vector diagram, respectively, for a centripetal ow compressor with two vector-adjusting stages, the Iirst vectoradjusting stage being positioned between two co-rotating compression stages and the second vector-adjusting stage single vector-adjusting stage.

FIG. 25 is a vector diagram for a Vclassical centripetal compressor having two contra-rotating rotors and a stationary contraprerot-ation input stage, with the energy contributions of the stages progressively decreasing as one progresses from the tirst stage toward the innermost stage.

FIGS. 26 and 27 are total head (pressure and kinetic energy) curves for the classical compressor and for two com-pressors having vector-adjusting stages.

FIG. 28 is a longitudinal sectional View of a complete gas turbine power plant disclosing a supersonic compressor with one vector-adjusting stage positioned between two co-rotating supersonic compression stages followed by one contra-rotating supersonic compression stage.

FIG. 29 is a transverse sectional View of the supersonic compressor illustrated in FIG. 28, said transverse section being taken along line 2929 shown in FIG. 2S. FIGS. 30A and 30B are vector diagrams for the supersonic compressor illustrated in FIGS. 28 and 29.

FIG. 31 is a vector diagram for a single rotation compressor and a contra-rotating compressor used for comparing the performances of the two types of compressors.

Consideration of Basic Problems Relating to Centrpelal Compression of Gases My U.S. Patent #2,712,895, entitled Centn'petal Subsonic Compressor, discloses a subsonic centripetal cornpressor having a stationary contra-prerotationjstage and two contra-rotating rotors rotatable at two equal `but opposite angular velocities. The successive compression stages of one rotor interleave the corresponding successive st-ages of the second rotor, and since the angular velocities of the two rotors are equal, the pressure heads of the successive stages decrease as a function of the squareof the `radiusof the respective stages. Accordingly, the outer compressor stage does the major part of energy conversion, and the last stage, or the inner stage', does the least amount of energy conversion. ,Such distribution of the energy conversion characteristics results in an over-all compression ratio which is limited by the` maximum Mach number which is permissible on the tirst stage. Accordingly, compression heads,

ft. lbs.

Patented June 26,1962

able with such compressors with the Mach number in the order of 0.9 in the iirst stage.

The centripetal flow, multi-stage contra-rotatable two rotor compressor disclosed in the U.S. Patent #2,712,895, has its velocity diagram illustrated in FIG. 12 in the above patent. For a better understanding of this invention, the same velocity vector diagram of FIG. 12 is also reproduced here in FIG. 25. If the iluid to be compressed is air, the ambient air enters the stationary prerotation stage of the compressor with an absolute velocity C (see FIG. 25) and leaves it with a velocity C1, which is the maximum absolute velocity of uid in the above compressor. The maximum peripheral velocity in the entire compressor is U1, which is the peripheral velocity of the outer rim of the first compression stage, and the maximum relative velocity is W1 at the entry into the irst stage. From then on, all Vectors progressively become smaller and the compressed air finally leaves the last stage with the lowest -absolute velocity C24 of the entire velocity vector system. It would be theoretically possible to convert the vector diagrams of FIG. 25 into a series of vector triangles equal to each other, in which case line 2500 and 2501 would be parallel to each other and also parallel to line 2504, which is the radius reference line for the vector diagram. Such conversion could be accomplished by making the peripheral velocities of all stages equal to the peripheral veloc- -ity U1. To obtain this, the inner stages would have to rotate at the progressively higher angular velocities, the angular velocities increasing from the outer stage toward the inner stage as a function of the mean radius of any given stage. If it were possible to achieve the above in actual practice with the aid of a practicable and inexpensive design, then all stages in a centripetal compressor would have equal energy conversions, The only practicable way of obtaining such peripheral velocity distribution would be by connecting the respective stages of the compressor to the respective stages of a free turbine, i.e., a turbine in which each stage is free to rotate at its own speed determined bythe forces acting on this stage. Structures of this type are disclosed in the Grifth U.S. Patent No. 2,391,779, in which the power plant comprises a plurality of concentric, disconnected, contrarotatable rings, each ring including one centripetal flow compression stage and one centrifugal flow turbine stage. Such attempts of obtaining equal energy conversions throughout the compressor are not practicably attainable due to the mechanical and iluid dynamic complexities.

Reverting once morerto FIG. 25, it is immediately apparent that the velocity vector diagram disclosed in FIG. 25 has only one critically high velocity, which is the relative velocity W1, at the entry into the first stage of the compressor, TheV local Mach number MWl at this point is equal to where a1 is the local speed of sound. This local Mach number determines the performance of the entire compressor of the above type and, what is most important, -it also determines the energy conversions of all the remaining stages of the compressor. The maximum value of MW1 may be in the order of .95 which is the maximum local Mach number that can be used in a subsonic compressor. (Theoretically the number is .999 If theflocal Mach number exceeds this maximum value, the compressor enters a supersonic region which will complicate the operation of a 'subsonic compressor utilizing subsonic airfoils. In centripetal compressors having straight blading, supersonic mode of compression begins the moment the local Mach number is equal to 1.0. In the axial flow machines there is the so-called transonic region, which corresponds to the mode of operation `at which the outer portion of the blade operates in the supersonic region, the root operates in the subsonic region, and there is a transonic region in the middle of the blade. Since in the centripetal flow vcompressor the blading is a straight blading, and the entire leading edge travels at the same velocity, the centripetal compressor blading operates either at subsonic or supersonic velocity, and there is no transonic mode of operation.

The inherent energy conversion distribution will remain basically the same as that illustrated in FIG. 25 and also in FIGS. 26 and 27 by curves 2600 and 2700. Curve 2660 illustrates the mechanical heads obtainable in each stage of the compressor, while curve 2700 illustrates the total head obtainable in the compressor. FIGS. 26 and 27 graphically illustrate what is also illustrated in FIG. 25, except that FIGS. 26 and 27 illustrate, so to speak, the end product, i.e., the mechanical heads (pressure and kinetic energies produced by any given compression stage) per stage (FIG. 26) and the total mechanical head (FIG. 27) attained by the compressor, when the velocity distribution is of the type illustrated in FIG. 25. The curves illustrated in FIGS. 26 and 27, are for MW,=.68.

For MW1=0.9, the values of the mechanical heads for all stages would be correspondingly higher. However, the curves of FIG. 26 would slope down just as rapidly as one would progress from the outer stages toward the innermost stage. The compressor disclosed in Patent #2,712,895 utilizing the velocity distribution illustrated in FIG. 25, will be called here, for simplifying all subsequent references, as a classical centripetal flow compressor. The meaning of the word classical here is that such compressor has two contra-rotatable rotors rotating at equal angular velocities, and the energy conversions of the stages decrease from the outermost stage toward the innermost stage as a function of the mean radius of any given stage, Le., the comprsor has the simplest geometry and vector diagram.

In the vector diagrams, as well as in the Eulerian relations in `column 3() and following, it is important to denne the vectorial directions of all peripheral vectors in tangential relation to the rotating and stationary stages. For his reason, vectorial signs are used only on tangential vectors and are omitted on all other vectors not tangent to the stages,` for simplifying the discussion.

In the classical compressor, from FIG. 25, one obtains the following relations, all velocities, except angular velocities, being vectors, counter-clockwise rotation being positive and clockwise rotation being negative:

iW11n+1=iC11n (in general for stage exit vectors) W112 is made equal to Cu2 in scalar terms by choice, for the first stage only, as shown in FIG. 25,

where MW1 is the local Mach number at the entry to the irst stage;

where W1il MWI: l

in scalar values;

K-l A022- 2 Mwa:

in scalar values;

W,12 is the peripheral component of W2; ft./sec.

W2 is the relative velocity at the exit from the first stage; ft./sec.

Cu2 is the peripheral component of C2; ft./sec.

C2 is the absolute exit velocity at the exit from the first compression stage; ft./sec.

nis the number of any stage.

w1 is the angular velocity of the first rotor of the compressor; ft./sec.

a2 is the angular velocity of the second rotor of the compressor; ft./sec,

A31 is the (static) sonic velocity in the gap before the lirst rotor; ft./sec. l

A32 is the (static) sonic velocity in the gap before the second rotor; ft./sec.

K is the ratio of specific heats in the given gap between two rotors;

C3 is the absolute entry velocity into the second rotor;

ft./sec.

In the above designation of the velocities, in the speciflcation as Well as in the drawings, the velocity changes that actually do occur in the interastage air gaps are neglected. This simplifies the nomenclature of the velocities and it is unnecessary to consider these velocity changes in the specification because they are not relevant insofar as the invention is concerned. These intergap velocities are always considered in designing the compressors and, therefore, neglecting of these velocity changes here is purely for simplifying the nomenclature as Well as the entire disclosure. However, it is not meant to convey the impression that these velocity changes are so insignificant that they can be neglected all the way around, even in the actual, practical design Work. This, of course, is not the case.

Another exception shouldY be also mentioned here and that is FIGURE 25. For a more accurate geometric presentation of all the velocity vectors and their angular relationships with respect to each other, the inter-gap velocity changes have been considered insofar as the construction of 'the geometric lines is concerned, but the nomenclature of the vectors, nevertheless, still remain the same as that used in the specification, i.e., the inter-gap velocity changes are neglected. Therefore, for example, the relative exit velocity from the first stage is designated as W2 and the absolute exit velocity is designated as C2. The entry velocities into the succeeding rotating stage are then designated as being W3 and C3 and it is also indicated that C25-C3, which is the equivalent to the designation in the algebraic form that the change in the inter-stage gap velocities is neglected. The above should be kept in mind in reading the specification as well as interpreting the meaning of the vectors in the vectorial diagrams.

In centripetal compressors, the absolute entry velocity into any rotor is greater than the absolute exit velocity from the immediately preceding rotor. This change is due to the free vortex law in the gap and has been neglected throughout the entire text, for reasons of simplicity.

Before proceeding with `a detailed description of several versions of the method disclosed here, it would be helpful to outline briefly the nature of these Versions.

(l) The rst version of the method improves the performance of the compressor by making Wu2 smaller than C2 which is accomplished, firstly, by increasing the angle of turning el of the 4iirst stage which is accomplished by increasing the camber and solidity of the first stage, and secondly, by increasing the angle of approach of the second stage. In this case, the vectorial relationship then becomes as follows:

The above relationship alsoV increases W3, which increases the loading of the second stage. Cu2 is made large enough to make W3 so large that the local Mach number MW3 is made equal to Mwl. The end result is that the compressor can create a static pressure head which is higher than the staticY pressure head obtainable with the classical compressor. No vector-adjusting stages are required in this version of the method. Higher compression ratio is obtained by making the vector diagram asymmetric. See FIGS. 2 through 4C.

(2) In the second version of lthe method, the conditions are identical to those in the first version but the peripheral velocity U3 is made equal to U1. Accordingly, the angular velocity m2 of the second, or inner rotor is made higher than the angular velocity w1 of the first, or outer rotor, the lirst rotor having a larger diameter than the diameter of the second rotor. The vectorial relationship, therefore, becomes as follows:

This compressor can create a total pressure head which is higher than the total pressure head obtainable with the compressor of the first version because of the increase in U3.

(3) In the third version one encounters the following conditions: The kinetic energy of the compressed iuid at the exit from the first stage, in the form of C2 and its peripheral component Cu2 (see FIG. 6) is so high that it cannot be effectively converted into pressure without exceeding the stipulated Mach number MW3 for the second stage. As will be explained later in a more detailed discussion of the third version of this invention, this vectorial relationship occurs ybecause the angle e2, which is the angle formed by C3 with the radial line 200, FIG. 2, is so large that the corresponding peripheral component of C3, which is (33 (in FIG. 2 it is designated as Cu2 because it is assumed here that C2f^-C3; this is a reasonably close approximation) is so large that when it is combined with the peripheral velocity vector U3 (representing the speed of travel of the outer rim of the second compression stage), the relative velocity vector W3 becomes excessively large. What is meant here by the expression that the velocity vector W3 becomes excessively large is that under such circumstances the local Mach number MW3 at the entry info the second stage will exceed 1.0 in subsonic or sonic compressors or will exceed a predetermined Mach number greater than 1 in a supersoniccompressor. Therefore, the only way that it would be possible under such circumstances to make the second stage work at a lower Mach number, would be by reducing all three vectors, i.e., C33, U3, and W3. Any reduction of U3 would reduce the total head produced not only by the second stage, but also by all the remaining inner stages of the compressor and therefore, an effective solution of the encountered problem cannot be looked for in the reduction of the magnitude of the peripheral velocity vector U3. If anything, U3 must remain constant and, for higher total head, which will be described more fully later, U3 should be made equal to U1 and, in some instances, it is made even higher than U1. This being the case, it, follows that there is only one vector which can be reduced in its magnitude without diminishing the total head, and that is the Cus vector. It should be realized here at once that the reduction of the C113 vector should be accomplished here without sacrifice in the magnitude of the absolute ow velocity vector C2 for the same reasons that were mentioned in connection with the vector U3, i.e., if the reduction in the magnitude of C23 were to be accomplished by reducing themagnitude of C2, such a reduction would be accomplished only at the expense of the compression ratio or total head of the compressor.

Stated differently, any reduction in C3 or C3 would at once produce a corresponding reduction in the over-all pressure head produced by the compressor. Accordingly, in this case, one encounters the following dilemma: one cannot reduce the magnitude of U3, one cannot reduce the magnitude of C3 or C3 without experiencing a corresponding loss in the total pressure head. In addition, it also should be stated here that no appreciable change can be obtained in the angle a3 because the magnitude of this angle is determined by the maximum turning angle e1 that can be obtained in the first compression stage without producing separations in the flow channels of the first stage. It may be mentioned here, if only parenthetically, that the maximum turning angle e1 of the first stage may be in the order of 70, and it would be more common to have this turning angle in the order of 60. This at once fixes the magnitude of a3, the design magnitude of which generally is between 35 and 45.

The third version of this method introduces a stationary vector-adjusting stage between the rst and the second compression stages. This vector-adjusting stage is capable of receiving C3 irrespective of its magnitude and direction and it is capable of changing the direction of this vector, without reducing its magnitude, to any desired extent by reducing angle a2 to any desired extent. As will be pointed out later more fully, this vector-adjusting stage also somewhat reduces the magnitude of vector C3(C3EC3) which is accomplished by obtaining a certain limited amount of diffusion in this vector-adjusting stage. Accordingly, the meaning of the term vector-adjusting stage as used in this specification means a stationary stage which receives the compressed ,fluid having one velocity vector, defined by its magnitude, direction of flow, such as C3, and discharges this compressed uid at Aa slightly lower velocity and different direction corresponding to a vector C1 in FIG. 6. The angle formed by C3 with the radial line 6G() in FIG. 6, which is angle a2, is larger than angle a4, formed by C4 with line 660. Accordingly, vectorially, one obtains the following rel-ationship:

The vectorial relationship and the Mach number relationship in the third variation of the invention, utilizing the vector-adjusting stage, FIG. 6, becomes as follows:

The significance of the first expression given above, i.e., that U1 may be equal, or approximately equal to or less than U3, is that the introduction of the vector-adjusting stage enables one to have much greater freedom in the design of all stages of the compressor following the vectoradjusting stage and one of the gained freedoms resides in the fact that it becomes possible to make U3 either equal to U1 or even larger than U1 by making w.1 w3. Accordingly, in the group of Expressions 6 it also includes the following (vectorially):

It is described more fully under the fourth Iversion of the method. From a Huid-dynamic point of view, the scalar magnitude of U3 in this case becomes limited only by Mwg, which should not exceed MW1 that may have any value assigned to it as a pre-determined characteristic parameter of the compressor. Accordingly, the introduction of the vector-adjusting stage enables one to produce 'a compressor with a constant Mach number in at least the first four compression stages. It will be pointed out later that, by introducing an additional vectoradjusting stage or stages, it is entirely practicable to produce a subsonic centripetal compressor in which the local Mach number remains approximately constant throughout the compressor. The constant Mach number throughout the compressor enables a centripetal flow contra-rotatable compressor to become by far the highest total head compressor for a given number lof stages. This high head, and the high compression ratio determined by it, is obtained with a mechanical structure and metals (conventional steel) which are much simpler and much cheaper than the structures of the axial flow or centrifugal flow compressors that would be necessary for obtaining the same compression ratio. The Aabove superiority is also obtained at much lower peripheral velocities.

(4) In the fourth variation, U3 is made larger than U1. This means that the peripheral velocity of the outermost compression stage of the second rotor is made larger than the peripheral velocity of the rst stage ducts w1 w3.

In this version, the vectorial and the Mach number relationships become `as follows (vectorially):

(5) In the fth version of the method, the vector and the Mach number relationships are as follows (vectorially) where U11, MW11 and Cun relate, or refer, to the nth stage `of the compressor; the nth stage being the rst stage of the second rotor after, or which follows, the last vector- 'adjusting stage. Therefore, it is one of the intermediate stages of the compressor. It is most likely that in this case the compressor will have more than one vector-adjusting stage.

This version of the method relates to the supersonic region of compressing an elastic uid.

Subsonic Compressor With Two Rotors Contra-Rotating at Two Opposite and Equal Angular Velocites but Having an Asymmetric Vector Diagram:

The type of compressor outlined in the `above title is the rst variation of the method. It is illustrated in FIGS. 1 and 2. FIG. l is the transverse cross-section of the compressor in a plane perpendicular to the axis of rotation and FiG. 2 is the vector diagram for this compressor.

Only three stages are illustrated in FIG. 1; they are the stationary contraprerotation stage 100, the rst compression stage 101 and the second compression stage 102. The stator stage 10) has a plurality of cambered airfoils 103. The uid enters the stator stage with a velocity C0 and leaves this stator stage with a velocity C1.

It will be assumed here, and throughout the specification, that the exit velocities from the stages are equal to the entry velocities into the succeeding stages, and therefore, C1 will be also considered as the rabsolute entry velocity of the fluid into the first compression stage (also, Ogm-C3; Cpe-C5, etc.). Actually, there is a slight increase in the absolute velocities `as the fluid crosses the interstage gaps; this increase need not be considered for the understanding of this invention, although, in actual design of the compressors, these changes `are taken into consideration.

Sufficient amount of contra-prerotation is imparted by the stationary contr'a-prerot-ation stage 100 so that the 9 direction of the relative velocity W1 coincides with the median low line 1104 of the flow channel in the rst compression stage at the point of entry of W1 into this stage, asis indicated in FIG. l.

The iirst compression stage has a plurality of carnbered airfoils 105, the camber being in the direction of rotation illustrated by an arrow, 106. The peripheral velocity ofthe outer rim is 'U1 While that of the inner rim is U2. The profile of the airfoils 105 was discussed earlier.

The second compression stage 102 is similar to the first stage. The airfoils 107 in general are of the same type as those in the iirst stage and the only difference between the lirst stage and the second, resides in the fact that the yangles ot stagger, y1 and y2 are dilerent, the angle of stagger 'y1 of the lirst stage being larger than y2. The variations of the airfoils from stage to stage is determined to a large extent by economic considerations which indicate the use of as few dilerent airfoils in the compressor as possible. However, lfor obtaining as high Reynolds numbers as possible, particularly in the lirst stages, it may be advisable to use a particular airioil having a longer chord than the chord of the airfoils used in the remaining stages of the compressor and this may lead one to the use of differently cambered airfoil and having different thickness in the rst compression stage than in other stages; for example, the lirst stage may have airfoils of the NACA-65(18)l0 type, while the succeeding stage may have an rairfoil of the NACA-65(16)09 type. The solidity of the respective stages is controlled primarily by the `angle of turning; the larger is the angle of turning, the larger is the solidity of the stage so long as the profile of the airfoils remains the same; on Vthe other hand, turning should be controlled by the camber of the blades instead of by variations of solidity.

The vectors for the second stage are C3, :W3 and U3; C3 is the absolute velocity of the fluid at the entry into the second stage, U3 is the peripheral velocity of the outer rim of the second stage, and W3 is the relative velocity of the fluid at the entry into the second stage. The vectors at the exit from the lirst stage are C2, W2 and U2. It is the position of these vectors with respect to the radius line 200 that differs from the position occupied by the same vectors wtih respect to the same radius line 2504 in the classical compressor that dilerentiates this compressor from the classical compressor.

As mentioned previously, in this case Wuz, illustrated in FIGS. 1 and 2, is made smaller than C112 which is achieved by increasing the angle of turning e1 of the first stage. It is the angle subtended by the relative velocity vectors W1 and W2. It is increased until Cu2 is sufliciently large to make the local Mach number MW3 at the entry into the second stage 102 equal to the local Mach number MW1 at the entry into the lirst stage 101. This process of adjusting the respective angles of turning of the respective compression stages continues throughout the compressor until the last stage, where the angle of turning is determined by the desired angle of exit of the absolute velocity Cn.

The vector relationship of the velocity component and of the local Mach numbers within the compressor in this case, is therefore of the type indicated by the group of Equations 3.

It is impossible to obtain a constant Mach number in `all of the compression stages in this case because w1=w2 and there is a decrease in the peripheral velocity of the stages as one progresses from the outer toward the innermost stage of the compressor. However, it is possible to obtain a constant entry Mach number in at least the rst three compression stages and, what is equally important, the decrease in the local Mach numbers in the remaining stages is not as rapid in this oase as in the classical compressor, with the result that all stages, without a single eX- ception, are more heavily yloaded and, therefore, produce higher total head per stage than the corresponding stages in the classical compressor.

rPhe above statement should be qualified by noting that the total head of the lirst stage of the compressor would be identical in both cases, so that the gain in the total compression is obtained in this case in the stages which follow the lirst stage. Also, the gain in the total head in this case is not as high as in the subsequent eases which will be described below. Nevertheless, even in this least favorable case, where no vector-adjusting stages are used, and w1=w2, this gain in the total mechanical head for the entire compressor may be in the order of 30%, this entire gain being achieved by la mere adjustment of the blade angles and Vectors, rather than by any additional structural means.

Subsonc Compressor with Two Rotors Contra-Rotating at Two Opposite Unequal Angular Velocies: w1 w2 The general configuration of this compressor is similar to t-hat illustrated in FIG. l and therefore, FIG. 1 may be used for its description.

The primary difference between the compressor illustrated in FIG. 1 and that contemplated here resides in the fact that in this case, as indicated in FIG. 3, w1 w2, and U3=U 1. Therefore, the stationary contra-prerotation stage is identical to that illustrated in FIG. 1. The lirst stage will diter from that illustrated in FIG. 1, in that its angle of turning e1 will be smaller than the same angle in FIG. l for the same Mach numbers. In FIGS. 2 and 3, the turning angles e1 are equal but the Mach numbers for FIG. 3 would `be higher than the corresponding Mach numbers in F IG. 2, and the Mach number would be much higher for W5 in FIG. 3 since W5 in FIG. 3 is much greater than W5 in FIG. 2. In both cases the stagger angle of the second stage will remain the same. The remaining odd-numbered stages will have a larger stagger angle because in this case, the components Wuz, WUG, Wulf), etc., will all be larger than in FIG. 2. The Vectorial relationship in this case -is illustrated in FIG. 3, and if this figure is compared with FIG. 2, then the following relationships are produced:

The above may be generalized by saying that all of the Cu components in FIG. 3 on the right and left sides of the radius line `300 have been increased, with the exception of a because of the increase in the peripheral velocity, U3, of the second rotor and over-all increase in the Mach number in FIG. 3. Because of the higher Mach number for all the stages in FIG. 3, it becomes necessary to transfer some of the kinetic energy from the first and second stages to the succeeding stages, and this has been accomplished by increasing Wu components, such as W115, on the left side of the radius lline 300.

This compressor, obviously, will produce a higher total head than the compressor of the rst version.

Szlbsom'c Compressor With Two Rotors Contra-Rotating at Two Opposite Angular Veloctes: w1 w2 FIGS. 4A-B-C illust-rate three possible changes in the vectors if the peripheral velocity U1 is increased to a peripheral velocity U1. In FIG. 4A the angles or positions of W1 and C1 with respect to the reference line 400 remain constant and therefore Athe only change that is produced by the increase in U1 is that illustrated by dotted lines in this figure. The original vector triangle U1W1C1 becomes larger and all of the vectors also become larger. This has been accomplished by extending W1 to W1', C1 to C1' until lthe distance between W1 and C1' is sufficiently large so as to accommodate the large vector U1. Tlhis increase in U1 to U1' and of U2 to U2' is also reected in the vertors representing the vectorial relationship at the exit from the first stage by the correspending increase in the magnitudes of C2 to C2 and W2 to W2' without any change in their angles. If the local Mach number at the entry into the first stage M1111, is less than 1.0, this solution of the .problem is entirely feasible so long as the increase in W1 is such as not to exceed the Mach number. Immediately after the .99 Mach number is exceeded, the solution becomes untenable and one must look for some other solutions which are described below in connection with FIGS. 4B and 4C.

In FIG. 4B the angular relationship of the vectors C1 and U1 remains the same, but the -angular position and magnitude of vector W1 has been changed by extending to the left that part of U1 which represents an increase in the peripheral velocity. Accordingly, the larger is the change from U1 to U1', the larger will be the change in the magnitude of W1 from W1 to W1', and of angle [l1 from [31 to [31', that W1 and W1 respectively, form with the radial lline 400. As to the magnitude of W1, as -long as W1 does not exceed the local Mach number of .99, this method is a feasible method. This solution, however, may not be feasible if the increase in the approach angle [81 is too large. The discussion of the optimum angles for the iirst compression stage will be given after concluding the discussion of FIG. 4C. U2, U2', [32, [32', W2, W2', are the corresponding vectors at the exit from hte first compression stage for two velocities U2 and U2'.

In FIG. 4C, an increase in U1 is accomplished by changing the magnitude as well as the angle of the absolute velocity C1. U1 is extended to the right and C1 is changed to C1'. From a purely vectorial point of View, this obviously is a feasible solution, and it is also feasible from the point of view of the loc-al Mach number. 'I'he local Mach number in this case will not be exceeded because the magnitude of W1 remains constant. The other changes that take place at the exit from the rst compression stage with the increase of U1 to U1 and the changing of angle a1 to a1' and C1 to C1' are as follows: a2 becomes u2', C2 becomes C2', and U2 becomes U2'.

However, other complications immediately arise in connection with FIG. 4C. For example, the angle of acceptance a1 changes quite rapidly and the magnitude of C111, which is the peripheral component of C1 (since line 40%) is passing through the center, or the axis of rotation, C111 is at right angles to line 400) is decreasing very rapidly as a1 approaches zero and, after this, the angle swings over to the right of the radial line 400. `C111 then changes its sign. In the notation used in this specification, Cu is positive when it lies to the right of line 406, and negative .when it lies to lthe left of line 400.

In general, total head f the rst stage is given by the following equation (this equation applies to any stage, single or contra-rotating, as long as proper velocity vectors are used in the equation):

ft2. lbs. lb.

where ft. lbs., lb.

AL is a mechanical head in dicated in FIG. 4C.

C2 is the absolute velocity of fluid at the exit from the first compression stage;

U1 is the peripheral Velocity at the entry into the lirst compression stage, or the peripheral velocity of the outer rim of the lirst compression stage, ft./sec.;

U2 is the peripheral velocity of the inner rim, or at the exit from the rst stage, ft/sec.

AL has a negative sign because it represents power input into the compressor, or `a shaft power supplied by an ex ternal source. This equation expresses the total conversion of mechanical energy by a compressor stage into pressure head energy and kinetic head energy. The obsolute value of this equation should be las high as possible in order to obtain maximum energy conversion from any given stage. The highest mechanical heads are obtained if both ter-ms in the bracket of Equation 10 can be made negative and numerically as large as possible. The peripheral component C111 has a positive value so long as C1 and C111 are to the left of the radius line 400, making the rst term, C1U1, negative since the vector U1 has 4a negative sign. The second term in the brackets, C112U1, is a positive product with a negative sign before it, because C112 is negative and U2 is also negative. Therefore, Equation l0 -for the vector diagrams illustrated in FIGS. 4A to 4C will have two negative terms on the right side of the equation so long as C1 is on the left side of the reference line 400 which produces large absolute values for the head AL. If now the increase of the peripheral Velocity U1 is accomplished at the expense of shortening of the vector C11, the total mechanical head is decreased (its absolute value) and if the velocity U1 is made equal to the peripheral component W111 of the relative velocity W1', then the peripheral velocity vector C111 becomes equal to zero. In this case, the total mechanical head converted by the rst stage is equal to:

-AL=C,12U2 (ll) A further increase of the peripheral velocity U1 results in transferring the position of vector C1 to the right side of radius line 460 and in setting up new angles of approach a1' (FIG. 4C) which causes the absolute velocity C1' to have a negative peripheral component C111'. lt can then be Seen that the total head is equal to:

-AL=%(+C,'U1f-CZU1') (11a) The capability of the stage to convert mechanical energy into compression is diminished relative to (11), although the total energy conversion is increased, due to the increased velocity, from U1 to U1'. Accordingly, it is undesirable to increase U1 by extending it in the manner in- If possible, it is desirable, therefore, lfor the peripheral component C111 to have a positive sign to the left of radius line 400 and numerically 4to be `as large as possible as long as its magnitude produces a proper angle of approach a1 and also the proper inflow angle [91. The optimum values of these angles as as follows:

1=any angle which is to the left of the radial line 400 so [S1=approximately 60 but preferably more than 60.

as to make C111 have a positive sign; 

